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Citation: Quintana, S.H.; Morales
Rojas, A.D.; Bedoya, I.D.
Experimental and Numerical
Evaluation of an HCCI Engine Fueled
with Biogas for Power Generation
under Sub-Atmospheric Conditions.
Energies 2023,16, 6267. https://
doi.org/10.3390/en16176267
Academic Editor: Amin Paykani
Received: 21 April 2023
Revised: 29 May 2023
Accepted: 1 June 2023
Published: 29 August 2023
Copyright: © 2023 by the authors.
Licensee MDPI, Basel, Switzerland.
This article is an open access article
distributed under the terms and
conditions of the Creative Commons
Attribution (CC BY) license (https://
creativecommons.org/licenses/by/
4.0/).
energies
Article
Experimental and Numerical Evaluation of an HCCI Engine
Fueled with Biogas for Power Generation under
Sub-Atmospheric Conditions
Sebastián H. Quintana 1,† , Andrés D. Morales Rojas 2,† and Iván D. Bedoya 1,*,†
1Grupo de Ciencia y Tecnología del Gas y Uso Racional de la Energía (GASURE), University of Antioquia,
Medellín 050010, Colombia; sebastian.herediaq@udea.edu.co
2Grupo de Investigación e Innovación en Energía (GIIEN), Institución Universitaria Pascual Bravo,
Medellín 050034, Colombia; andres.morales@pascualbravo.edu.co
*Correspondence: ivan.bedoya@udea.edu.co
† These authors contributed equally to this work.
Abstract:
Energy transition to renewable sources and more efficient technologies is needed for
sustainable development. Although this transition is expected to take a longer time in developing
countries, strategies that have been widely explored by the international academic community,
such as advanced combustion modes and microgeneration, could be implemented more easily.
However, the implementation of these well-known strategies in developing countries requires in-
depth research because of the specific technical, environmental, social, and economic conditions.
The present research relies on the use of biogas-fueled HCCI engines for power generation under
sub-atmospheric conditions provided by high altitudes above sea level in Colombia. A small air-
cooled commercial Diesel engine was modified to run in HCCI combustion mode by controlling the
air–biogas mixture temperature using an electric heater at a high speed of 1800 revolutions per minute.
An experimental setup was implemented to measure and control the most important experimental
variables, such as engine speed, biogas flow rate, intake temperature, crank angle degree, intake
pressure, NOx emissions, and in-cylinder pressure. High intake temperature requirements of around
320
◦
C were needed to achieve stable HCCI combustion; the maximum net indicated mean effective
pressure (IMEPn) was around 1.5 bar, and the highest net indicated efficiency was close to 32%.
Higher intake pressures and the addition of ozone to the intake mixture were numerically studied as
ways to reduce the intake temperature requirements for stable HCCI combustion and improve engine
performance. These strategies were studied using a one-zone model along with detailed chemical
kinetics, and the model was adjusted using the experimental results. The simulation results showed
that the addition of 500 ppm of ozone could reduce the intake temperature requirements by around
50
◦
C. The experimental and numerical results achieved in this research are important for the design
and implementation of HCCI engines running biogas for microgeneration systems in developing
countries which exhibit more difficult conditions for HCCI combustion implementation.
Keywords:
homogeneous charge compression ignition (HCCI); power generation; biogas; numerical
simulation; single-zone model
1. Introduction
The World Energy Outlook by the International Energy Agency (IEA) recently stated
that increasing energy efficiency and increasing the use of renewable energy are two of the
most important strategies for achieving a scenario of sustainability development with net
zero emissions [
1
]. Biogas is a renewable fuel that can be produced from organic waste,
and has great potential for use in power generation, transportation, and the synthesis of
fine chemicals and industrial products [
2
]. The global demand for biogas was around
35 billion cubic meter equivalents (bcme) in 2021, and is expected to increase to 138 bcme
Energies 2023,16, 6267. https://doi.org/10.3390/en16176267 https://www.mdpi.com/journal/energies
Energies 2023,16, 6267 2 of 21
in 2050, according to the Net Zero Emissions scenario [
1
]. A traditional way of using
biogas is power generation via internal combustion engine, with a minimum of methane
(CH
4
) concentration between 21% and 30% (volumetric basis) and hydrogen sulfide (H
2
S)
concentration below 500 ppm [
3
]. The most efficient way to use biogas for power generation
is through diesel engines in dual-fuel combustion mode, achieving diesel substitution close
to 90% in stationary applications. The efficiency of end-use technologies and their capacity
to operate solely on biogas are the factors that define the role of this fuel in expected energy
scenarios [
4
]; additionally, the implementation of the cleanest and most advanced kinds of
combustion is necessary because of the high costs of emissions after-treatment systems.
Homogeneous charge compression ignition (HCCI) engines have been studied for five
decades as an alternative to vehicle powertrains because of their potential for simultaneous
reduction of nitrogen oxides (NO
x
) and particulate matter (PM) without emissions after-
treatment systems. HCCI combustion is usually implemented through commercial engine
modification by removing the OEM ignition control and implementing strategies that con-
trol the intake charge conditions, such as intake pre-heating, negative valve overlap, and
boost pressures. HCCI is a combustion mode mainly controlled by chemical kinetics, with
little effect of turbulence and higher heat release rates when compared with compression
ignition engines and spark-ignited engines. Due to its rapid heat release rate, lean equiva-
lence ratios or highly diluted mixtures are required in order to obtain safe operation under
HCCI combustion, leading to in-cylinder temperatures below 1800 K, which in turn leads
to ultra-low NOxand PM emissions [5]. In HCCI engines, the beginning of combustion is
controlled by the charge auto-ignition temperature; therefore, fuels with high auto-ignition
resistance require higher intake temperatures to achieve stable and safe combustion. Biogas
is manly composed of methane (CH
4
) and carbon dioxide (CO
2
). Methane exhibits higher
auto-ignition temperatures when compared with other hydrocarbons, and carbon dioxide
reduces the average charge temperature during compression stroke because of its lower
specific heats ratio when compared with air; this combined effect leads to high intake
temperature requirements for biogas-fueled HCCI engines.
One of the most important drawbacks of the implementation of HCCI combustion is
the lowered net indicated mean effective pressure (IMEPn), which is usually around 5 bar
for naturally aspirated engines operating at sea level with liquid fuels. Any factor that
reduces the intake air density tends to decrease the achievable IMEPn in HCCI engines.
The implementation of HCCI combustion for fuels with higher auto-ignition temperatures
and operating at high altitudes above sea level requires substantial research efforts in order
to determine the potential of HCCI engines for power generation under these adverse
conditions. The most important previous research on this specific field, to the best of the
authors’ knowledge, is presented below.
Blizman et al. [
6
] evaluated a Caterpillar 3116 6.6-L diesel engine converted to operate
in HCCI mode fueled with both natural gas and simulated landfill gas. The engine com-
pression ratio and speed were 16.7:1 and 1800 rpm, respectively. Operation with simulated
landfill gas was evaluated at equivalence ratios of 0.35, 0.36, 0.38, and 0.41, with intake tem-
peratures ranging between 160 and 185 ◦C and boosted intake air around 1.4 bar absolute
pressure. Operating with simulated landfill gas as fuel, a 33% Brake Thermal Efficiency
(BTE) was achieved while producing about 26 kW of electrical power, compared to the 35%
and 29 kW, respectively, achieved with natural gas. NOx and unburned hydrocarbon (HC)
emissions were maintained below 10 ppm.
Bedoya et al. [
7
] studied a Volkswagen TDI 1.9-L diesel engine in HCCI mode and
fueled with simulated biogas (60% CH
4
and 40% CO
2
in a volumetric basis) with a modified
compression ratio of 17:1 and running at 1800 rpm. Baseline tests were carried out at
equivalence ratios of 0.25, 0.30, and 0.40, with absolute intake pressures of 2 and 2.2 bar and
intake temperatures between 473 and 483 K. Excessive cycle-to-cycle variation (normalized
standard deviation of IMEP
g
greater than 10%) was observed at the lowest equivalence
ratio, while unacceptable ringing intensity (around 14 MW/m
2
) was observed at the highest
equivalence ratio. In later research, Bedoya et al. [
8
] used two strategies at low equivalence
Energies 2023,16, 6267 3 of 21
ratio to improve engine performance, specifically, oxygen enrichment of the intake charge
and gasoline pilot port injection. At high equivalence ratio, delayed CA50 after top dead
center was used to reduce excessive ringing intensities. Oxygen enrichment did not allow
for stabilized combustion at the low load limit. Gasoline pilot port injection decreased
cycle-to-cycle variation, HC, and carbon monoxide (CO) emissions. The delayed CA50
allowed led to a reduction in ringing intensity. In summary, a biogas-fueled HCCI engine
operating with stable combustion between 0.2 and 0.5 of equivalence ratios was successfully
implemented, allowing indicated efficiencies above 40%, a maximum gross indicated mean
effective pressure (IMEPg) of 8.5 bar, and NOx emissions below 20 ppm.
Kozarac et al. [
9
] performed a numerical evaluation of the effect of internal and
external charge recirculation on the combustion variables of an HCCI engine fueled with
biogas, and investigated the potential of these strategies to reduce the requirement for
a high intake temperature. Their study revealed that the evaluated strategies did not
significantly reduce the required intake temperature for a fixed combustion timing. On
the other hand, they observed that high internal charge recirculation, which was achieved
using negative valve overlap (NVO), led to unstable combustion.
The studies mentioned above suggest that stable HCCI combustion with biogas is
feasible within a narrow range of equivalence ratios, and that multiple strategies are
necessary to reduce both cyclic variability at low loads and ringing intensities at high
loads. A potentially useful strategy for a power generation system based on biogas-fueled
HCCI engines is the addition of ozone to the intake charge. This strategy requires less
maintenance compared to charge heating, and demands less energy to operate. Ozone
addition tends to reduce the ignition delay as the mixture becomes leaner, as it promotes O
radical formation, leading to a greater concentration of OH and H radicals [
10
]. Nishida
and Tachibana [
11
] experimentally and numerically evaluated the effect of adding ozone to
the charge of an HCCI engine fueled with natural gas, finding that an ozone concentration
of 700 ppm could potentially lead to the same thermal efficiency with respect to diesel
engine operation. In additional, they were able to achieve low NOx emissions. The authors
further highlighted the possibility of using ozone addition as an ignition timing control
strategy during HCCI combustion.
The present research focused on studying biogas HCCI combustion in a microgen-
eration system at 1500 m above sea level, which provides sub-atmospheric conditions
(85 kPa of intake pressure). A small commercial diesel engine was modified to allow biogas
auto-ignition close to the top dead center by controlling the intake temperature and equiva-
lence ratios. Results on the intake temperature requirements and the main performance
parameters for the practical operating range for biogas HCCI combustion are reported in
this paper. These experimental results are important for the implementation of microgen-
eration systems running biogas with HCCI combustion, as to the best of our knowledge
there is a lack of research on the practical limits for this kind of combustion operating at
sub-atmospheric conditions when using renewable fuels such as biogas. Additionally, our
research aimed to perform a numerical study of boosted intake pressures and addition of
ozone to the intake charge as strategies for reducing the required intake temperature for
stable combustion and increasing the engine’s power output.
2. Materials and Methods
2.1. Experimental Setup
The research engine used in this study was a modified two-cylinder four-stroke air-
cooled and naturally aspirated direct-injection diesel engine. An electric motor was used
to keep the engine speed constant at 1800 rpm. The experiments were conducted in
Medellín, Colombia (1500 m above sea level). Table 1shows the most important technical
engine characteristics.
Energies 2023,16, 6267 4 of 21
Table 1. Technical engine specifications.
Designation
Lombardini 25LD425/2 Direct-Injection Four
Stroke Two-Cylinder Air-Cooled
Diesel Engine
Charge aspiration Naturally aspirated
Displacement 851 cm3
Original compression ratio 19:1
Modified compression ratio 17.4:1
Bore ×stroke (mm) 85 ×75
Connecting rod length (mm) 117.5
Valves (intake, exhaust) 1, 1
Intake Valve Open, IVO (CAD ATDC) 350
Intake Valve Close, IVC (CAD ATDC) −115
Exhaust Valve Open, EVO (CAD ATDC) 95
Exhaust Valve Close, EVC (CAD ATDC) −325
Original rated power 14 kW @ 3600 rpm
Original maximum torque 40.5 Nm @ 2400 rpm
Original combustion chamber Bowl
Modified combustion chamber Flat head
Engine modifications included the replacement of the original bowl-type pistons with
flat pistons. This change was carried out to reduce the inlet temperature requirements
by lowering the effective area for heat transfer. Additionally, the compression ratio was
reduced from 19:1 to 17.4:1. Fuel was supplied through a fuel injection system in the
intake manifold using a set of calibrated orifices to measure the biogas flow rate. A
piezoelectric transducer (Kistler 6125C) coupled to a charge amplifier (Kistler 5064B) was
used to capture the in-cylinder pressure in a cylinder. A Kistler SCP 2853A120 system was
used for pressure signal conditioning, while the pressure measurements were triggered
by a crankshaft encoder with a resolution of 10 samples per crank angle degree (Nord
8.5820H40.1024). The air flow rate was estimated by measuring the intake pressure, intake
temperature, and engine speed. An electronic pressure gauge (Wika A10) was used for
measuring intake pressure. Inlet temperature was measured with K-type thermocouples.
CA50, calculated from the in-cylinder pressure data, was monitored in real-time and served
as a feedback parameter for the intake temperature using an open-loop control, which was
controlled by adjusting an intake air heater (Silvania threaded inline heater 1-1
/
4
00
NPT
6 kW).
The engine was fueled using a mix of commercial compressed natural gas (CNG)
and CO
2
to simulate a biogas composition with values of 60% CNG and 40% CO
2
on a
volumetric basis. Table 2summarizes the important properties of the natural gas used in
the experimental procedure. Table 3shows the most relevant data on the accuracy and
uncertainty of the instruments used during the experiments.
Table 2. Fuel properties.
Property Natural Gas Biogas
Low heating value (MJ/kg) 47.26 18.51
Simplified chemical
composition C1.166H4.271 O0.0015N0.0032 -
Stoichiometric air fuel
ratio (AFR) 16.13 6.31
Lower Woobe index
(kWh/Nm3)16.39 6.31
Methane number [12] 71.74 -
Energies 2023,16, 6267 5 of 21
Table 3. Accuracy and uncertainty associated with measurements.
Measurement Accuracy Uncertainty
Fuel flow rate (mg/s) ±3 6.5%
Air flow rate (g/s) ±0.1 5%
Temperature (◦C) ±0.2 2%
Intake manifold pressure (bar)
±0.01 2%
In-cylinder pressure (bar) ±0.0008 0.3%
Crank angle position ±0.1 4%
Engine speed (rpm) ±10 3%
All signals were recorded on a personal computer. The acquisition system recorded
data at a rate of 250 kHz with a resolution of 16 bits. The experimental setup is shown in
Figure 1.
Figure 1.
Schematic diagram of the experimental setup; T: temperature measurement, P: pressure
measurement.
2.1.1. Experimental Procedure
Two equivalence ratios (low and high value) were evaluated using two intake temper-
atures. The factorial experimental design was conducted using the methodology described
by Montgomery [
13
]. The levels and factors evaluated in the biogas tests are shown in
Table 4, which were replicated twice.
Table 4. Factorial experimental design for biogas experiments.
Factor Level Description Level Designation
Intake temperature 1 325 ◦C
2 330 ◦C
Equivalence ratio 1 0.28
2 0.42
Engine speed 1 1800 rpm
A total of 100 cycles were recorded for each operating point.
Energies 2023,16, 6267 6 of 21
2.1.2. Rate of Heat Release Calculations
Heat release rate (HRR) was calculated from the acquired in-cylinder pressure curves
using a zero-dimensional heat release model. Consequently, the main combustion pa-
rameters were extracted from the heat release and in-cylinder pressure curves. HRR was
calculated as follows [14]:
dQchem
dθ=γ
γ−1pdV
dθ+1
γ−1Vdp
dθ+dQw
dθ+hcr dmcr
dθ(1)
where
Qchem
is the chemical energy released by the fuel (J),
γ
is the specific heat ratio,
p
is the in-cylinder pressure (Pa),
V
is the instantaneous cylinder volume (m
3
),
Qw
is the
gas-surrounding heat exchange (J),
hcr
and
mcr
are the enthalpy (J/kg) and mass (kg) of the
flow in the piston-cylinder-ring crevices, respectively, and θis the crank angle.
Heat transfer was calculated using the correlation of Chang et al. [15]:
hg=αsL−0.2 p0.8T−0.73v0.8 (2)
v=C1Sp+C2
6
VdTI VC
pIVCVIVC
(p−pmot )(3)
where
hg
(W/m
2
.K) is the heat transfer coefficient,
αs
is a scaling factor of the heat transfer
coefficient (2.5 in this research),
L
(m) is the instantaneous chamber height,
p
(kPa) is
the in-cylinder pressure,
T
(K) is the gas temperature calculated from the ideal gas state
equation,
v
(m/s) is the combustion induced velocity,
Sp
(m/s) is the mean piston speed,
Vd
(m
3
) is the cylinder displacement,
pIVC
(kPa),
TIVC
(K), and
VIVC
(m
3
) are the gas
conditions at IVC, and
pmot
(kPa) is the motoring pressure (without combustion). Moreover,
C1
is a dimensionless constant with a value of 2.28, while
C2
(m/s.K) is a constant which
takes a value of 0 during the compression period and 0.00324 during the combustion and
expansion period.
Flow in the piston cylinder ring crevices was modelled as an orifice flow [16]:
˙
mcr =CdAprγ
RT 2
γ+1γ+1
2(γ−1)(4)
When:
patm
p≤2
γ+1γ
γ−1(5)
Otherwise:
˙
mcr =CdArpr1
RT patm
p1
γ
2γ
γ−1
1−patm
pγ−1
γ
1
2
(6)
In Equations
(4)
–
(6)
,
˙
mcr
is the mass flow rate in the piston–cylinder–ring crevices
(kg/s),
Cd
is the discharge coefficient (0.7 in this research),
Ar
is the area of the orifice
connecting the combustion chamber and the crankcase (which is considered equal to the
passage area caused by ring gaps, m
2
),
p
is the in-cylinder pressure (kPa),
patm
is the
pressure in the crankcase (which is considered equal to atmospheric pressure in kPa), γis
the specific heat ratio,
R
is the gas constant (kJ/kg.K), and
T
is the mass temperature (K),
which is assumed to be equal to the wall temperature.
The mass contained inside the cylinder immediately after of the intake valve closure
consists of air, fuel, and residual gas from previous cycles. The residual gas fraction was
calculated using the expression proposed by Ortiz-Soto et al. [17].
Energies 2023,16, 6267 7 of 21
The start and end of combustion were calculated using the 10% (CA10) and 90%
(CA90) of mass burned (xb), respectively, which was calculated as follows:
xb=RdQchem dθ
mfLHVf
(7)
where
mf
is the mass of fuel inducted per cycle and
LHVf
is the low heating value of
the fuel.
Finally, the normalized standard deviation of IMEPg was used to evaluate the com-
bustion stability of HCCI, which is calculated using the following expression:
NStd.Dev.o f IME Pg=Std.Dev.o f IMEPg
IMEPg−IMEPg,motored
∗100 (8)
where
Std
.
Dev
.
o f I MEPg
is the standard deviation of measured
IMEPg
for 100 cycles
and
IMEPg,motored
is the
IMEPg
in motored conditions. The maximum value for acceptable
stability was set to 3% for the normalized standard deviation of
IMEPg
, in accordance with
the values suggested in the literature [18].
2.2. Numerical Methodology: Single-Zone Model Approach
A CHEMKIN zero-dimensional detailed kinetic model was used to model HCCI
auto-ignition to evaluate the effect of increasing intake pressure and ozone addition on
combustion phasing and intake temperature requirements. This is a lumped model that
ignores spatial variations in the combustion chamber and treats heat loss as a distributed
heat transfer rate that is proportional to the temperature difference between the average gas
temperature and the time-averaged wall temperature. Although the model oversimplifies
the combustion chamber conditions during combustion, because the heat release is a global
non-propagating auto-ignition process a zero-dimensional model can reasonably capture
the start of combustion and the heat release of the core temperature zone. Predictions for
start of combustion and for NOx depend on the peak temperature of the core gases inside
the cylinder. Therefore, the model is expected to provide reasonable accuracy for estimating
optimal operating conditions and certain combustion performance parameters [19].
To reduce over-prediction of the pressure during compression stroke and near to the
TDC, the compression ratio was adjusted to match the experimental motored pressure and
numerical motored pressure. In these simulations, a compression ratio of 15.3:1 was used,
while the bore, stroke, and connecting rod were identical to the experiments (see Table 1).
The heat transfer to the cylinder walls was modeled using the correlation developed
by Woschni to estimate the average cylinder gas speed and Nusselt number:
w=C11 +C12
vswirl
SpSp+C2
VdTi
piVi
(p−pmotored)(9)
Nu =aRebPrc(10)
The values used for the constants were
C11 =
2.28,
C12 =
0.308,
vswirl
Sp=
2.0,
C2=
0.054,
a=0.0234, b=0.8, and c=0.
The optimal operating conditions associated with ozone addition or increased intake
pressure were defined from the ignition timing, where the crank angle of ignition was
defined as the crank angle for the peak molar fraction of H
2
O
2
, which usually coincides
with the crank angle for 10% of cumulative heat release (CA10) in simulations [
20
]. The
crank angle interval selected for the peak molar fraction of H
2
O
2
to ensure an optimal
operating condition was between 3 CAD before top dead center (bTDC) and 3 CAD after
top dead center (aTDC), as high heat release rates below this interval have usually been
obtained in previous research, leading to excessive ringing intensity, while above this
interval high cycle-to-cycle variations have been obtained, leading to misfire [7,21].
Energies 2023,16, 6267 8 of 21
Two gas phase mechanisms were used to analyse the effects of the kinetic model
on the operating condition results. The mechanisms used were GRI 3.0 (51 species and
325 reactions) and San Diego (63 species and 307 reactions, including nitrogen chemistry),
version 2014-10-04. A comparison shows that both mechanisms have 46 species and
144 reactions in common (the same reactants and products), and that 139 of these reactions
show differences in at least one of the Arrhenius parameters.
The effect of ozone on auto-ignition and combustion was studied using the ozone
kinetic submechanism used by Halter et al. to study the effect of ozone on methane
combustion [
22
]. This submechanism was coupled to both of the detailed mechanisms used
in this research.
Computations were performed during the closed-valve period between IVC and EVO
(see Table 1). Two different ozone concentrations (100 and 250 ppm) were used to evaluate
this strategy on the intake temperature required to reach the auto-ignition condition near
the TDC. In addition, three different pressures at the IVC were used (1, 1.5, and 2 bar) in
order to compare the effect of ozone concentration with another well-known strategy for
improving engine performance and the energy requirement in the intake.
2.3. Wall Temperature Estimation
Wall temperature was estimated using polytropic coefficient and Woschni correla-
tions [
23
]. Polytropic coefficients were calculated using the in-cylinder pressure, as follows:
n=log∆p
log∆V(11)
where nis the polytropic coeffcient.
The compression polytropic coefficient was calculated between 90 and 80 CAD bTDC,
while the expansion polytropic coefficient was calculated between 60 and 90 CAD aTDC.
Figure 2shows the values obtained for both coefficients at different equivalence ratios. It
can be observed that the polytropic coefficients show few variations with the equivalence
ratio. For this reason, the compression, and expansion polytropic coefficients were assumed
as constants.
After polytropic coefficient was calculated, the heat transferred between the in-cylinder
gas and the wall was estimated as follows:
δQw=n−γ
1−γpδV(12)
After calculating the gas–surroundings heat exchange, the wall temperature was
estimated using Newton’s law of cooling:
Tw=T−δQw
hgA(13)
where Twis the wall temperature and Ais the heat transfer area.
Heat transfer coefficient was calculated as follows:
h=Nuk
B(14)
where the Nusselt number
Nu
is calculated using Equation
(10)
(with the Reynolds number
calculated using Woschni’s correlation for the average cylinder gas speed in Equation
(3)
),
k
is the thermal conductivity of the gas, and
B
is the bore cylinder. Due to the lean equivalence
ratio, the thermal conductivity and dynamic viscosity of the charge are assumed to be equal
to those of air.
Energies 2023,16, 6267 9 of 21
Figure 2.
Compression and expansion polytropic coefficients at different equivalence ratio conditions
for the engine used in the research.
3. Results and Discussions
3.1. Experimental Results
Figures 3and 4show the respective CA10 and CA50 related to the evaluated equiv-
alence ratios and intake charge temperatures required for stable combustion. An intake
temperature of 325
◦
C was necessary to allow for HCCI combustion at the tested conditions,
and previous research performed by Bedoya et al. [
7
] showed that intake temperatures
around 200
◦
C allowed for stable HCCI combustion at sea level and with boosted inlet
pressures close to 2 bar. For the same intake temperature, it can be seen that both CA10
and CA50 were delayed when the equivalence ratio was increased. This trend is associated
with a higher concentration of CO
2
in the intake mixture with higher equivalence ratios.
However, the delayed combustion effect was reduced at higher intake temperature due to
the higher in-cylinder temperature.
Figure 3. Crank angle for 10% cumulative heat release; Pint =0.84 bar.
Energies 2023,16, 6267 10 of 21
Figure 4. Crank angle for 50% cumulative heat release; Pint =0.84 bar.
Figures 5and 6show the respective in-cylinder pressure and gross heat release rates
at the different intake temperatures and equivalence ratios evaluated. As expected, the
pressure peak was increased, and was closer to TDC for higher intake temperatures, and
this trend was more obvious for the higher tested equivalence ratios. Similar trends were
observed for the peaks of the gross heat release rates. The maximum pressure reached for
the conditions evaluated in this research were well below the maximum pressure reached
in previous research for similar compression ratios [
7
–
9
] due to the lower intake pressure
and for the higher required intake temperatures, which together lead to greatly reduced
power output.
Figure 5. In-cylinder pressure; Pint =0.84 bar.
Energies 2023,16, 6267 11 of 21
Figure 6. Gross heat release rates; Pint =0.84 bar.
Figure 7shows the normalized standard deviation of IMEPg and Figure 8shows the
percentage of residual gases, with both figures related by the equivalence ratio. When the
equivalence ratio was increased, combustion was delayed due to higher concentration of
CO
2
in the charge (see Figures 3and 4). However, cycle-to-cycle variations were signifi-
cantly reduced, as can be seen in Figure 7. It can be observed that at an equivalence ratio
of 0.28 the engine operated with higher combustion instability, especially at the intake
temperature of 325
◦
C. This trend is explained by the higher percentage of residual gases
presented at lower equivalence ratios, as can be seen in Figure 8. Higher charge dilution
leads to increased cyclic variability for heat release rates and in-cylinder pressure traces [
9
].
In Figure 8, it can be observed that the estimated residual gases inside the cylinder from the
previous cycle are greater that the expected values for compression ignition engines [
14
],
which is explained by the high crank angle interval at which the intake and exhaust valves
remain open in the exhaust and intake strokes (see IVC and EVO in Table 1). The greater val-
ues of residual gases for the equivalence ratio of 0.28 compared with the ratio of 0.42 explain
the lowered pressure peak when the intake temperature increases (see Figure 5).
The high intake temperature requirement to operate an HCCI engine fueled with
biogas at sub-atmospheric conditions has a strong effect on power output because of the
reduced volumetric efficiency at these conditions. Figures 9and 10 show the IMEPn and
indicated efficiency, respectively, under the tested conditions. The highest IMEpn was
around 1.5 bar for an equivalence ratio of 0.42 and intake temperature of 325
◦
C. Previous
research performed on biogas-fueled HCCI engines at sea level by Bedoya et al. [
7
] reported
an IMEPg of about 5 bar for boosted intake pressure close to 2 bar. As expected, higher
intake temperatures led to reduced IMEPn because of the reduced charge density. The
higher net indicated efficiencies were around 32% for this study, and were lower compared
with the aforementioned research on biogas HCCI combustion [
7
], in which gross indicated
efficiencies close to 45% were reported.
Energies 2023,16, 6267 12 of 21
Figure 7. Coefficient of variation of IMEPn; Pint =0.84 bar.
Figure 8. Percentage of estimated residual gases inside the cylinder; Pint =0.84 bar.
Energies 2023,16, 6267 13 of 21
Figure 9.
IMEPn at different conditions of equivalence ratio and intake temperature; P
int =
0.84 bar.
Figure 10.
Indicated efficiency at different conditions of equivalence ratio and intake temperature;
Pint =0.84 bar.
Energies 2023,16, 6267 14 of 21
Figure 11 shows the NOx emissions in relation to the evaluated conditions. As
expected, NOx emissions were increased with higher equivalence ratios; however, the
concentrations remained ultra-low, which is a feature of HCCI combustion. Similar results
on NOx emissions were reported in previous research by Bedoya et al. [7].
Figure 11.
Emissions of NOx at different conditions of equivalence ratio and intake temperature;
Pint =0.84 bar.
3.2. Numerical Results
The analysed experimental results have shown that operating an HCCI engine fueled
with biogas at sub-atmospheric conditions leads to lowered IMEP and indicated efficiencies
as compared with previous research performed at sea level. For this reason, the effect of
increasing intake pressure and intake ozone addition on intake temperature requirements
was evaluated using a single-zone model with detailed chemical kinetics.
To ensure that the numerical results had adequate physical correspondence with the
experimental results expected when the same strategies were applied, a set of simulations
was run using the single-zone approach.
As mentioned above, the combustion timing was defined as the crank angle for the
peak molar fraction of H
2
O
2
. The experimental IVC conditions were used to run simulations
under the single-zone approach. The charge composition was defined according to the esti-
mated experimental equivalence ratios and residual gas fractions (RGF).
Figures 12 and 13
show the simulated and experimental ignition timings. The results are shown for the GRI
3.0 and San Diego mechanisms (both the original mechanism and the version including the
ozone kinetics submechanism from Halter et al. [
22
]). In Figure 12, it can be seen that the
GRI 3.0 mechanism shows better agreement with experimental data for the condition when
combustion is delayed. However, when combustion is highly delayed the model predicts
ignition timings closer to TDC. This could be associated with the effects of turbulence
and charge stratification for ignition timings greater than 2 CAD aTDC, which are not
completely accounted for in the simplified single-zone model. On the other hand, the two
San Diego mechanisms show lower agreement with the experimental data for delayed
conditions, and better agreement for ignition timings near the TDC compared to GRI 3.0. In
Energies 2023,16, 6267 15 of 21
Figure 13, it can be observed that inclusion of the ozone submechanism has a slight effect
on the simulated results, showing slight combustion advancement. Thus, it can be assumed
that any effect on combustion advancement when ozone is added to the charge and the
ozone submechanism is used is strictly related to the kinetic effect of this species.
Figure 12. Simulated and experimental ignition timing using original mechanisms. Pint =0.84 bar.
Figure 13.
Simulated and experimental ignition timing using mechanisms with ozone kinetic
sub.mechanism; Pint =0.84 bar.
Energies 2023,16, 6267 16 of 21
The results presented in Figures 12 and 13 show that the single-zone approach can be
used to estimate the IVC conditions that lead to safe ignition timings (i.e., without high
cycle-to-cycle variations or excessive ringing intensity) even under delayed combustion
conditions. However, for ignition timings greater than 2 CAD aTDC the effect of turbu-
lence and charge stratification on this process become significant, and the results are no
longer suitable.
3.3. Effect of Boosted Intake Pressure on Intake Temperature
Figures 14–16 show the effect of increasing the intake pressure on the intake temper-
ature required to reach safe operating conditions. It can be observed that is possible to
reduce the intake temperature by up to 25
◦
C for the same equivalence ratio and ignition
timing when the intake pressure is increased around to 2.4 times. This effect has been
reported in previous papers [24–26].
The important information conveyed by Figures 14–16 is the temperature required
for safe operation, which is shown in the bottom graph of each figure. This temperature
was always observed to be 50 and 60
◦
C higher than the temperature calculated during
the experiments at the IVC. The figures show that the lowest temperature requirements
are associated with medium-lean equivalence ratios and delayed ignition timings, while
high-lean and advanced ignition timings require the highest temperatures. In the case of
the engine used in the research, high temperatures above 290
◦
C were required to achieve
a safe operating condition in the zone of delayed ignition timing, which could adversely
affect engine efficiency and lead to derating compared to diesel operation, this being a
drawback of using HCCI combustion in microgeneration with biogas or other biofuels with
high resistance to auto-ignition. Additionally, these high temperatures could impact the
lubrication system due to heating of the in-cylinder surfaces during the intake stroke, which
could modify the lubricant properties and increase the wear on engine parts, potentially
reducing engine lifespan and increasing the frequency of required maintenance.
Figure 14.
IVC (
top
) and intake (
bottom
) temperature maps for safe operating condition;
Pint =1 bar
and 1800 rpm.
Energies 2023,16, 6267 17 of 21
Figure 15.
IVC (
top
) and intake (
bottom
) temperature maps for safe operating condition;
Pint =1.5 bar and 1800 rpm.
Figure 16.
IVC (
top
) and intake (
bottom
) temperature maps for safe operating condition;
Pint =2 bar
and 1800 rpm.
Energies 2023,16, 6267 18 of 21
3.4. Effect of Ozone Addition on Intake Temperature
Figures 17 and 18 show the effect of ozone addition on the intake temperature required
to reach safe operating condition for additions of 100 and 500 ppm, respectively. For ozone
addition, a molar balance was used to ensure the same C/O ratio as the baseline condition
(i.e., entry of only biogas and air). It can be observed in the Figure 17 that addition of
only 100 ppm of ozone reduces the intake temperature required to reach safe operating
condition by around 50
◦
C. Compared to the baseline conditions, the intake temperature is
around 280
◦
C for an ultra-lean equivalence ratio and delayed ignition timing, while for a
lean equivalence ratio and advanced ignition timing the intake temperature is around to
300
◦
C. This represents a significant improvement in engine power output and derating
compared to the baseline operating conditions.
When ozone addition is increased to 500 ppm (as shown in Figure 18), the required
intake temperature to reach safe operating conditions decreases to around 250
◦
C for an
ultra-lean equivalence ratio and delayed ignition timing. Similarly, for a lean equivalence
ratio and advanced ignition timing, the intake temperature decreases to around 270 ◦C.
Regarding to the boosted intake pressure strategy, the ozone addition strategy shows
better performance in reducing the required intake temperature. The reduction is between
20
◦
C and 50
◦
C with ozone addition of 100 ppm and 500 ppm, respectively, compared to
an intake pressure of 2 bar at the same equivalence ratio and ignition timing. With respect
to the baseline, the ozone addition strategy appears to be a better option for reducing the
intake temperature of a high-speed HCCI engine fueled with biogas for power generation
at sub-atmospheric conditions.
Figure 17.
IVC (
top
) and intake (
bottom
) temperature maps for safe operating condition;
Pint =0.84 bar, ozone addition=100 ppm, and 1800 rpm.
Energies 2023,16, 6267 19 of 21
Figure 18.
IVC (
top
) and intake (
bottom
) temperature maps for safe operating condition;
Pint =0.84 bar, ozone addition=500 ppm, and 1800 rpm.
4. Conclusions
In this research, an air-cooled naturally aspirated Diesel engine operating at high
engine speed was conditioned to run in HCCI mode while being fueled with biogas for
power generation. Experimental tests were conducted under sub-atmospheric conditions
(1500 m.a.s.l.). The four conditions studied here (high engine speed, biogas as fuel, an
air-cooled system, and sub-atmospheric conditions) represent technical challenges, as these
conditions negatively affect the auto-ignition process. To reach HCCI combustion based
on combustion duration and NO
x
concentrations, high intake temperatures above 320
◦
C
were required and highly diluted charges (close to 12%) were obtained. These conditions
led to low IMEPn (around 1.5 bar) and low indicated efficiency (close to 32%), which are
below the values reported in previous studies on biogas-fueled HCCI combustion. To
explore potential improvements in the performance of the tested engine, two strategies to
reduce intake temperature were explored using a zero-dimensional detailed kinetic model,
specifically, boosted intake pressure and ozone addition. Boosting the intake pressure up to
2 bar gauge can reduced the intake temperature to around 300
◦
C for centered combustion,
while addition of 500 ppm ozone to the charge allows the intake temperature to be reduced
to around 250
◦
C. These results show that ozone addition is a better alternative to increase
the IMEPn and indicated efficiency in HCCI engines fueled with biogas thanks to the
lower intake temperature required for stable combustion. The results achieved in this
study are important for the design and implementatoin of biogas-fueled HCCI engines for
microgeneration systems at high altitudes above sea level, which is a common scenario in
developing countries. Future research on other strategies that can improve the performance
of HCCI engines running biogas will be addressed by the authors in future studies involving
more complete numerical approaches that allow the estimation of regulated emissions and
combustion instabilities.
Energies 2023,16, 6267 20 of 21
Author Contributions:
Conceptualization, S.H.Q. and I.D.B.; Methodology, S.H.Q. and A.D.M.R.;
Software, S.H.Q.; Investigation, S.H.Q.; Writing—original draft, S.H.Q.; Writing—review & editing,
S.H.Q., A.D.M.R. and I.D.B.; Supervision, I.D.B. All authors have read and agreed to the published
version of the manuscript.
Funding:
The authors gratefully acknowledge the financial support provided by the Colombia
Scientific Program within the framework of the Ecosistema Científico (Contract No. FP44842-218-
2018). The authors acknowledge the financial support provided by the University of Antioquia
through the research project identified by the code PRG 2017-16268.
Data Availability Statement: Not applicable.
Conflicts of Interest:
The authors declare no conflict of interest. The funders had no role in the design
of the study; in the collection, analyses, or interpretation of data; in the writing of the manuscript; or
in the decision to publish the results.
Abbreviations
The following abbreviations are used in this manuscript:
aTDC After top dead centre
bcme Billion cubic metres equivalent
bTDC Before top dead centre
int Intake
rpm Revolutions per minute
AFR Air fuel ratio
BTE Brake Thermal Efficiency
CA10 Crank angle of 10% of cumulative heat release
CA50 Crank angle of 50% of cumulative heat release
CAD Crank angle degree
CNG Compressed natural gas
CO Carbon monoxide
CO2Carbon dioxide
EVO Exhaust valve opening
GHG Greenhouse gases
H Hydrogen
HC unburned hydrocarbons
HCCI Homogeneous Charge Compression Ignition
HRR Heat release rate
ICE Internal combustion engine
IMEP Indicated Mean Effective Pressure
IMEPg Gross indicated mean effective pressure
IMEPn Net indicated mean effective pressure
IVC Intake valve closure
LHV Low heating value
NOxNitrogen oxides
NVO Negative valve overlap
PM Particulate matter
PV Photovoltaic
RGF Residual gas fraction
TDC Top dead center
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